Ratio Shift Control System And Method For A Multiple-Ratio Automatic  Transmission

ABSTRACT

A control system and method for controlling a multiple gear ratio automatic transmission in a powertrain for an automatic transmission having pressure activated friction torque elements to effect gear ratio upshifts. The friction torque elements are synchronously engaged and released during a torque phase of an upshift event as torque from a powertrain source is increased while allowing the off-going friction elements to slip, followed by an inertia phase during which torque from a powertrain source is modulated. A perceptible transmission output torque reduction during an upshift is avoided.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is a continuation of application Ser. No. 12/858,468,filed Aug. 18, 2010, which is a continuation-in-part of application Ser.No. 12/693,086, now abandoned, filed Jan. 25, 2010, which is assigned tothe assignee of the present application and the disclosures of which areincorporated in their entirety by reference herein.

BACKGROUND

The invention relates to a multiple-ratio transmission mechanism in apowertrain for an automotive vehicle and to a control strategy forachieving smooth engagement and release of friction torque establishingelements during a transmission upshift event.

In a geared automatic transmission in an automotive vehicle powertrainhaving an engine or other torque source, a ratio change may be made froma so-called low ratio to a so-called higher ratio when a friction torqueestablishing element, such as a clutch or brake, is engaged insynchronism with disengagement of a companion friction torqueestablishing element. This is referred to as a ratio upshift. Thefriction torque establishing elements involved in the upshift may bereferred to as an oncoming clutch or brake and an off-going clutch orbrake. The upshift event is characterized by a preparatory phase, atorque phase and an inertia phase as the vehicle accelerates from astanding start.

In a conventional automatic transmission in a vehicle powertrain, theoncoming clutch torque capacity is controlled to increase from a lowvalue during the torque phase. Simultaneous engagement of one clutch orbrake and release of another results in a momentary activation of twotorque flow paths through the gearing, causing a gear tie-up in whichtransmission output shaft torque decreases momentarily. This conditionmay be referred to as a “torque hole”. It occurs before the off-goingclutch totally disengages.

Friction elements, such as disc clutches, band brakes and disc brakes,typically are actuated hydraulically under the control of a transmissioncontrol module, which disengages an off-going friction clutch or brakewhile simultaneously engaging an oncoming friction clutch or brakeduring an upshift in order to lower speed ratio. For purposes of thepresent description of the invention, the clutch and the brake will bereferred to as friction elements.

During the preparatory phase, an automatic transmission control reducesoff-going friction element torque capacity to prepare it for release asan actuator for the oncoming friction element is adjusted to prepare forits engagement. During the torque phase, the controller increasesoncoming friction element torque capacity, which causes torquetransmitted through the off-going friction element to drop quickly dueto the transient gear tie up.

As torque is transmitted through the off-going friction elementdeceases, the automatic transmission output shaft torque drops, whichcauses the so-called torque hole. This is perceived by a vehicleoccupant as an unpleasant shift shock. The inertia phase begins when theoff-going clutch is released with no significant torque capacity.

SUMMARY

The invention comprises a transmission ratio control system and methodthat eliminates or reduces a so-called torque hole during upshifting oftransmission gearing of a step ratio automatic transmission. Theautomatic transmission, for example, can be either a layshafttransmission with two torque input friction elements between a torquesource and the transmission gearing, or a step ratio automatictransmission with planetary gearing, wherein a ratio change in thegearing during an upshifting event is effected by engaging one torqueinput friction element for the gearing and simultaneously disengaginganother torque input friction element for the gearing. For purposes ofdescribing the present invention, reference will be made to a lay-shafttype transmission.

The invention includes a strategy for execution of control algorithmsthat will achieve a desired output shaft torque profile that will avoidsignificant output shaft torque disturbances.

In the case of a powertrain with an internal combustion engine, torqueinput to the automatic transmission is increased during the torque phaseof the shifting event. This is achieved by engine throttle control,spark timing adjustment for the engine (torque source), intake andexhaust valve timing control for the engine or by other means, such asby using auxiliary electric motor torque, based on an open loop control,a closed loop control, or a combination of both using engine speed,off-going and oncoming clutch slip speed measurements, and clutchactuator position measurements.

According to one aspect of the invention, a software-based controller isprovided to self-calibrate a level of oncoming clutch torque capacityusing algorithms in the form of algebraic equations whereby a desiredoutput shaft torque profile is achieved while the off-going clutch slipsduring the torque phase in a controlled manner.

According to another aspect of the invention, the desired output shafttorque profile is achieved for a chosen off-going clutch torquecapacity.

The invention, in executing the foregoing control features, may decouplecontrol of engine torque or the input shaft torque from an oncomingclutch torque control during the torque phase, while the off-goingclutch slips, and to achieve a desired off-going clutch slip based on aclosed loop control of input shaft torque or engine torque. The end ofthe torque phase is determined based on torque level transmitted throughthe off-going clutch.

According to a further aspect of the invention, governing algebraicequations are used to determine a level of the oncoming clutch torquecapacity to achieve a seamless transition from the torque phase to theinertia phase. This involves a self-calibration of a level of oncomingclutch torque capacity during the inertia phase to achieve a desiredoutput shaft torque level.

In one embodiment of the invention, the off-going friction element isallowed to slip during the torque phase of a shift event as slip of theoncoming friction element is controlled.

According to another aspect of the invention, input torque may beincreased during the torque phase, and the change in torque may be usedin a determination of torque capacity of the off-going friction elementduring the torque phase. Torque of the torque source is reduced duringthe inertia phase and then restored, at least partially, after theinertia phase.

According to another aspect of the invention, control of the oncomingclutch torque control is decoupled from engine or input shaft controlduring the torque phase, and a desired off-going clutch slip is achievedbased on a closed-loop control of input shaft torque (e.g., enginetorque).

According to another aspect of the invention, the end of the torquephase is determined based on the torque level transmitted through theoff-going clutch.

According to another aspect of the invention, a target level of oncomingclutch torque capacity is determined using governing equations toachieve a seamless output shaft torque transition from the torque phaseto the inertia phase.

According to another aspect of the invention, a target level of oncomingclutch torque capacity is determined during the inertia phase usinggoverning equations to achieve a desired output shaft torque level.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic illustration of a layshaft transmission in a firstgear or low gear operating mode, which includes tandem torque inputclutches that are selectively and alternately engaged and released.

FIG. 1 a is a schematic illustration of the gearing arrangement of FIG.1 wherein the elements of the gearing are conditioned for high or secondgear operating mode.

FIG. 2 is a schematic representation of a planetary type transmissionthat is capable of embodying the invention wherein the elements of theplanetary gearing are conditioned for a low or first gear operatingmode.

FIG. 2 a is a schematic representation corresponding to FIG. 2 whereinthe elements are conditioned for a second or high gear operating mode.

FIG. 2 b is a schematic representation of another planetary transmissionthat is capable of embodying the invention.

FIG. 3 is a time plot for a synchronous clutch-to-clutch upshift controlcharacterized by a so-called torque hole at the output shaft.

FIG. 4 is a time plot corresponding to FIG. 3 for the synchronousupshift control of the present invention wherein the tome hole iseliminated.

FIG. 5 is a flowchart showing the control strategy of the synchronousupshift control of the present invention when the off-going clutch isslipping.

FIG. 6 is a flowchart showing an alternate control strategy for anon-synchronous upshift when the off-going clutch is slipping.

FIG. 7 is a flowchart showing another alternate control strategy for anon-synchronous upshift when the off-going clutch is slipping.

PARTICULAR DESCRIPTION OF AN EMBODIMENT OF THE INVENTION

As required, detailed embodiments of the present invention are disclosedherein; however, it is to be understood that the disclosed embodimentsare merely exemplary of the invention that may be embodied in variousand alternative forms. The figures are not necessarily to scale; somefeatures may be exaggerated or minimized to show details of particularcomponents. Therefore, specific structural and functional detailsdisclosed herein are not to be interpreted as limiting, but merely as arepresentative basis for teaching one skilled in the art to variouslyemploy the present invention.

FIG. 1 shows a schematic form of a lay-shaft transmission capable ofembodying the invention together with a schematic representation of thetransmission components involved in gear ratio changes.

Numeral 10 represents a power input shaft drivably connected to torquesource 12. Input shaft 10 drives a clutch housing 14, which carriestorque input driving discs 16 situated in inter-digital relationshipwith respect to driven discs 18 and 20. A fluid pressure actuator orelectro-mechanical actuator of any known design is used to selectivelyengage driven discs 18 and 20 with respect to driving discs 16. Discs 20are connected to a central torque input shaft 22 and discs 18 areconnected to torque input sleeve shaft 24. Although only one disc 18 andonly one disc 20 are shown in the schematic view of FIGS. 1 and 1 a,several discs in a friction disc assembly may be used.

Drive gear elements 26 and 28 are connected drivably to the sleeve shaft24. Gear element 26 has a smaller pitch diameter than gear element 28.

Central power input shaft 22 is drivably connected to drive gear element30, gear element 32 and gear element 34, which have decreasing pitchdiameters.

When driving clutch discs 20 are engaged, driving torque is distributedthrough engaged clutch discs 20 to the gear elements 30, 32 and 34.Clutch discs 20 and 18 are part of the clutch structure that may bereferred to as a tandem or dual clutch 36.

When clutch discs 18 are engaged by the tandem clutch 36, torque fromthe torque source is distributed directly to torque input gears 26 and28.

The layshaft transmission of FIG. 1 has two countershafts, shown at 38and 40. Countershaft 38 supports rotatably a third ratio countershaftgear element 40, a fourth ratio countershaft gear element 42 and areverse countershaft gear element 44. A torque transfer gear element 46is directly connected to the countershaft 38.

Countershaft 40 rotatably supports countershaft gear elements 48, 50 and52, which have progressively decreasing pitch diameters. Countershaftgear element 48 is a first ratio gear element, countershaft gear element50 is a fifth ratio gear element and countershaft gear element 52 is asixth ratio gear element.

Countershaft gear elements 54 and 56 also are rotatably supported bycountershaft 40. Gear element 54 drivably engages gear element 26 duringsecond ratio operation. Countershaft gear element 56 drivably engages areverse drive pinion (not shown), which in turn drivably engages reversegear element 44 during reverse drive operation. Gear element 46connected to countershaft 38 is drivably connected to gear element 58,which is drivably connected to countershaft 40, for example, throughtorque transfer gearing (not shown in FIG. 1). The countershafts and thecentral shaft 22 actually are not in the same plane, so torque transfergearing and the reverse drive pinions are not illustrated in theschematic illustration of FIG. 1.

Gear 58 is connected drivably to torque output gear 60, which isdrivably connected to vehicle traction wheels.

During first gear ratio operation, gear 48 is connected drivably throughsynchronizer clutch 62 to countershaft 40, and clutch 36 engages discs20 as discs 18 are disengaged. At that time, second ratio synchronizerclutch 64 drivably engages gear element 54 to precondition gear element54 for second ratio operation. Power then is delivered from the torquesource through clutch discs 20 to central shaft 22 so that torque isdelivered from gear 34, to countershaft 40 and engaged gears 58 and 60.

An upshift is made from the first gear ratio to the second gear ratio bydisengaging clutch discs 20 and engaging clutch disc 18 for the tandemclutch. To make a smooth transition from the first gear ratio to thesecond gear ratio, discs 18 are engaged as discs 20 are slowlydisengaged to allow for clutch slip. At this time, third ratiosynchronizer clutch 66 is engaged thereby connecting countershaft gearelement 40 to countershaft 38. This preselects third ratio while thetransmission operates in the second ratio. An upshift to the third ratiois achieved by tandem clutch 36 as clutch discs 20 are engaged andclutch discs 18 are disengaged. At this time, the fourth ratiosynchronizer clutch 68 is engaged to preselect the fourth ratio. Anupshift from the third gear ratio to the fourth gear ratio then isachieved by disengaging clutch discs 20 and engaging clutch discs 18. Atthis time, fifth gear ratio is preselected by engaging synchronizerclutch 70. An upshift to the fifth ratio then is achieved by engagingfriction discs 20 and disengaging friction discs 18. At this time, thesixth ratio is preselected by engaging synchronizer clutch 72.

An upshift to the sixth ratio is achieved by again trading engagement ofthe discs for the tandem clutch 36. Clutch discs 20 are disengaged asclutch discs 18 are engaged.

Reverse drive is obtained by disengaging the forward drive synchronizerclutch and engaging reverse drive synchronizer clutch 74. Reversedriving torque then is delivered through sleeve shaft 24, gear 26, gearelement 54 and gear element 56, reverse drive pinion gearing,countershaft 38 and torque transfer gear elements 46 and 58.

If the torque source is an internal combustion engine, the upshiftcontrols would include a microprocessor 75, which may be of conventionaldesign, an electronic engine control 77, including an engine fuel andspark retard controller, and a transmission control module 83.

The microprocessor 75 receives, when the torque source is an engine,input signals such as driver desired input torque (Te_des) input speed(Ne), driver-selected ratio range (PRNDL), transmission input speed(Ninput), engine throttle position (Tp) if the torque source is athrottle-controlled engine, and transmission output speed (Noutput). Theinput signals are received by random access memory (RAM) from data inputports. A central processor unit (CPU) receives the input signals thatare stored in RAM and uses the information fetched from RAM to executealgorithms that define control strategies stored in ROM. Output signalsare delivered from signal output ports to the controllers 77 and 83.Actuating pressure for the clutches is supplied by pump 85 driven byengine 12 or by an electro-magnetic force actuator.

FIG. 1 a shows the gearing configuration during operation of thetransmission in second gear ratio, which is the upshifted ratio. Whenthe transmission operates in the second ratio, torque is delivered, aspreviously mentioned, to sleeve shaft 24 and through a second gear set,which comprises gear 26, gear element 54 and transfer gears 58 and 60.This gearing may be referred to as the second gear set. The gearingpreviously described with respect to FIG. 1 for first gear operationhereafter may be referred to as the first gear set.

FIGS. 2 and 2 a show a schematic representation of a planetary typetransmission that may embody the present invention. A torque source maybe an engine 76 that drives a ring gear 80 of a simple planetary gearunit 82, which has a sun gear 84 and a planetary carrier 86. Ahydrokinetic torque converter may be included in the transmission if adesign objective requires it. It is shown at 78 in FIGS. 2 and 2 a withphantom dotted lines since some designs capable of using the inventiondo no need a torque converter. If a torque converter is included, theconverter turbine torque would be the input torque. The torque convertercould be deleted if it is not needed. Carrier 86 supports planetarypinions that engage ring gear 80 and sun gear 84. The output torque fromthe carrier drives sun gear 88 of a compound planetary gear set 90.Compound planetary pinions 92 and 94 supported on a common carrier 96engage respectively ring gear 90 and sun gear 88. The ring gear isconnected to the output shaft 98.

During low gear ratio operation, friction brake 100 is disengaged. Brake100 may be referred to as clutch #1. This corresponds to tandem clutch36 of FIGS. 1 and 1 a when clutch discs 18 are released or disengaged.Brake 102 in FIG. 2, which is engaged in low speed ratio operation,corresponds to tandem clutch 36 shown in FIGS. 1 and 1 a when clutchdiscs 20 are engaged. Clutch #2 in FIG. 2 (brake 102) provides areaction point for the carrier 96. Sun gear, shown at 104, whichdrivably engages with compound planetary pinion 92, merely idles duringlow speed ratio operation.

When the gearing of FIGS. 2 and 2 a is operating in the second ratio,sun gear 104 is anchored by brake 100 so that the ring gear for compoundplanetary gear unit 92 is driven at an increased rate relative to thecarrier speed of the simple planetary gear set 82.

For purposes of this description, it will be assumed that if thepowertrain has no hydrokinetic torque converter, torque input to thetransmission will be referred to as engine torque (Te). If thepowertrain has a torque converter, the engine torque would be replacedby converter turbine torque.

FIG. 2 b shows an example of another planetary step-ratio automatictransmission that may embody the invention. It comprises an enginedriven torque input shaft 11 and a transmission input shaft 13. Atransmission output shaft 15 delivers torque to transmission torqueoutput gearing 17. A torque converter may be disposed between enginedriven torque input shaft 11 and a transmission input shaft 13, as shownat 19. A torque converter impeller 11 is in fluid flow relationship withrespect to turbine 13. A stator 15 is disposed between the flow inletsection of impeller 11 and the flow exit section of turbine 13.

In the example of a planetary transmission shown in FIG. 2 b, there arethree simply planetary gear units 21, 23 and 25. Output torque isdelivered from the carrier 27 to the torque output gearing. Carrier 27is connected to the ring gear for gear unit 25 and to output shaft 15.An overrunning coupling 29 anchors the carrier 31 of planetary gear unit25 against rotation in one direction, but free wheeling motion isprovided in the opposite direction. During reverse and during low ratiooperation, carrier 31 is braked by coupling 33 against the transmissionhousing 35. During forward drive operation, the sun gear for gear unit21 is anchored to the housing through forward drive coupling 37.

During intermediate ratio operation, the sun gear for gear unit 25 isanchored to the housing 35 by intermediate coupling 39.

During direct drive, the transmission input shaft 13 is clutched bydirect coupling 41 to input shaft 13, thus establishing a one-to-onedriving ratio through the planetary gearing. Overdrive coupling 43, whenengaged, directly connects the carrier for gear unit 25 and the ringgear for gear unit 23 to the input shaft 13. FIG. 1 a shows an engine 12which acts as a source of torque for the transmission. If thetransmission has a torque converter, engine speed will equal speed ofconverter impeller 22 and transmission input speed would equal converterturbine speed.

FIG. 3 shows a strategy for a typical known upshift event from a lowgear configuration (i.e., high torque ratio) to a high gearconfiguration (i.e., low torque ratio) when the engine has a constantthrottle setting, in accordance with a conventional upshift controlmethod for a lay-shaft transmission of the type shown in FIGS. 1 and 1a. This strategy of the invention would apply also to a transmissionsuch as the compound planetary transmission of FIGS. 2 and 2 a and theplanetary transmission of FIG. 2 b.

The shift event is divided into a preparatory phase, a torque phase, andan inertia phase. During the preparatory phase, torque capacity ofclutch 20, which is the off-going clutch, is reduced, as shown at 86, toprepare for its release. However, enough clutch torque capacity ismaintained at 88 to only allows a small incipient slip near the end ofthe preparatory phase, as shown by the small separation between thedotted input torque line 106 and OGC line 86. Transmission controller 82adjusts an actuator piston for clutch 18 (clutch #2), which is referredto as the oncoming clutch, to prepare for its engagement. At that point,the oncoming clutch 18, in a synchronous upshift event, is yet to carrysignificant torque capacity.

During the torque phase of the control shown in FIG. 3, off-going clutchcapacity is further reduced, as shown at 91, while the controller 82increases oncoming clutch torque capacity, as shown at 93. Engine speedand input shaft speed are the same if the transmission has no torqueconverter between the engine and the clutch 36. However, as will beexplained subsequently in a discussion of FIG. 4, off-going clutchtorque capacity may be controlled to induce a small target level slip at91, which allows engine speed 95 to be higher than the speed of shaft22. When the off-going clutch slips, off-going clutch torque 91, orfrictional torque generated by slipping, drives shaft 22, seen in FIGS.1 and 1 a and the downstream gear elements (gearset #1), all the way tothe output shaft. Increasing oncoming clutch torque 93 starts balancingtorque distributed from the engine and reduces the off-going clutchtorque capacity requirement at 91. Thus, the off-going clutch and theoncoming clutch work together to maintain off-going clutch target levelslip as the off-going clutch torque decreases as shown at 91.

During the torque phase of the shift characteristic shown in FIG. 3, anincrease in oncoming clutch torque capacity (clutch #2 capacity) reducesnet torque flow through the off-going clutch when the off-going clutchremains engaged. Thus, the output shaft torque drops significantly, asshown at 97, creating a so-called torque hole. A large torque hole canbe perceived by a vehicle occupant as a sluggish powertrain performanceor an unpleasant shift shock.

The inertia phase begins when the off-going clutch capacity is reducedto a non-significant level, as shown at 98. Oncoming clutch (clutch #2)carries enough torque capacity, as shown at 100, to pull down enginespeed, as shown at 102, closer to that of the speed of shaft #2, asindicated at 104.

FIG. 3 shows reduced input torque during the inertia phase, as shown at106. This is typically due to engine spark timing control, which iscommon practice in the conventional shift control method, to enable theoncoming clutch to engage within a target shift duration withoutexcessive torque capacity.

The shift event is completed, as shown in FIG. 3, when clutch #2 (theoncoming clutch) is engaged. The input shaft then is securely coupled toshaft 24, seen in FIG. 1, thereby matching engine speed 102 to shaftspeed 104. The engine torque reduction at 106 is removed at 108 and theoutput shaft torque returns to the level that corresponds to an enginetorque level during the high gear configuration.

In contrast to the upshift characteristics shown in FIG. 3, FIG. 4 showsthe upshift characteristics of an embodiment of the upshift controlmethod of the invention. During the preparatory phase, the controller 83reduces the torque capacity of the off-going clutch (discs 20) toprepare for its release, as shown at 110. The controller also adjuststhe actuator piston for clutch 18 (the oncoming clutch) to prepare forits engagement.

During the torque phase, the controller 83 increases oncoming clutchtorque capacity, as shown at 112, to prepare for its engagement. Inputtorque is increased, as shown at 114, while allowing clutch discs 20 toslip at a controlled level. Slipping the off-going clutch discs 20causes input speed to be slightly greater, as shown at 124, than theshaft speed, shown at 116. This is true for a transmission having aslipping off-going clutch, but it is not true for a transmission with alocked off-going clutch.

When the off-going clutch 20 slips, its torque capacity or frictionaltorque is transmitted to shaft 22. Thus, the transmission controller canactively manage torque level that drives the gears coupled to thegearing connected to shaft 22 by adjusting the off-going clutch torquecapacity 118. Similarly, when the oncoming clutch slips during thetorque phase, its torque capacity, shown at 112, is transmitted to shaft24, which drives the gearing (gearset #2) connected to shaft 24. Thus,when both the off-going clutch (OGC) and the oncoming clutch (OCC) slipduring the torque phase, output shaft torque τos can be mathematicallydescribed as:

τ_(os) =G _(on)τ_(on) +G _(off)τ_(off),  Eq. (1)

where τ_(on) is OCC torque capacity, τ_(off) is OGC torque capacity,G_(off) is gear ratio for low gear operation and G_(on) is gear ratiofor high gear operation. Equation (1) can be rearranged as:

$\begin{matrix}{\tau_{on} = \frac{\tau_{os} - {G_{off}\tau_{off}}}{G_{on}}} & {{Eq}.\mspace{14mu} (2)}\end{matrix}$

Rewriting τ_(os) as τ_(os,des), Eq. (2) can be expressed as:

$\begin{matrix}{{\tau_{on} = \frac{\tau_{{os},{des}} - {G_{off}\tau_{off}}}{G_{on}}},} & {{Eq}.\mspace{14mu} (3)}\end{matrix}$

where τ_(os,des) is a desired output shaft torque. The governingequation (3) of the present invention provides a systematic means toself-calibrate a level of OCC torque capacity τ_(on) for achieving adesired output torque profile τ_(os,des) while OGC slips during thetorque phase. More specifically, torque profile τ_(os,des) can bespecified to smoothly transition output shaft torque 120 before andafter the torque phase, from point 71 to point 73 and after point 73,thereby eliminating or reducing the torque hole. OGC torque capacityτ_(off) can be estimated and actively adjusted based on OGC actuatorposition or clamping force. Thus, for a given τ_(off), Eq. (2) specifiesa level of OCC torque capacity τ_(on) (112) required for achieving adesired output shaft torque 120.

During the torque phase, powertrain controller 75 and engine controller77 control engine torque 114 or input shaft torque in order to maintainOGC slip at a desired level. This can be achieved, for example, byadjusting engine torque 114 using a closed-loop throttle control, valvetiming control or fuel control or engine spark timing control based onOGC slip measurements independently from OCC and OGC torque control in aseparate control loop or background loop, for the controller.

The transmission controller 83 (FIG. 1) could maintain enough OGC torquecapacity during the torque phase without allowing OGC to slip. In thiscase, OGC still transmits a part of engine torque 114 to shaft #1 (22).

Output shaft torque is described as:

τ_(os) =G _(off)τ_(in)+(G _(on) −G _(off))τ_(on),  Eq. (4)

where input shaft torque τ_(in) can be equated to input torque τ_(e)(when the transmission has no torque converter). Replacing τ_(os) with adesired torque profile τ_(os,des), Eq. (4) can be rearranged as:

$\begin{matrix}{\tau_{on} = {{\frac{\tau_{{os},{des}} - {G_{off}\tau_{e}}}{G_{on} - G_{off}}\mspace{14mu} {or}\mspace{14mu} \tau_{e}} = {\frac{\tau_{{os},{des}} - {\left( {G_{on} - G_{off}} \right)\tau_{on}}}{G_{off}}.}}} & {{Eq}.\mspace{14mu} (5)}\end{matrix}$

Torque variables τ_(os) and τ_(e) can be represented as:

τ_(os,des)=τ_(os) ₀ −Δτ_(os) and τ_(e)=τ_(e) ₀ +Δτ_(e),  Eq. (6)

where τ_(os0) and τ_(e0) are the output shaft torque and engine torqueat the beginning of the torque phase, respectively. Δτ_(os) and Δτ_(e)represent the change in output shaft torque and engine torque,respectively, at the elapsed time At after the torque phase begins.Substituting Eq. (6) into Eq. (5) yields:

$\begin{matrix}{\tau_{on} = {\frac{{\Delta\tau}_{{os},{des}} + {G_{off}{\Delta\tau}_{e}}}{G_{off} - G_{on}}.}} & {{Eq}.\mspace{14mu} (7)}\end{matrix}$

OCC torque τ_(on) can be written as:

τ_(on)=τ_(on) ₀ +Δτ_(on),  Eq. (8)

where τ_(on0) is the OCC torque capacity at the beginning of the torquephase and Δτ_(on) is the change in OCC torque at Δt. Substituting Eq.(8) into Eq. (7) results in:

$\begin{matrix}{{{\Delta \; \tau_{on}} = \frac{{\Delta\tau}_{{os},{des}} - {G_{off}{\Delta\tau}_{off}}}{G_{on}}},} & {{Eq}.\mspace{14mu} (9)}\end{matrix}$

where Δτ_(off)≡τ_(e)−Δτ_(on). (Note that Eq. (9) takes the same form asEq. (3), which is the governing equation for slipping OGC.)

The governing equations (5),(7) and (9) provide a systematic means toself-calibrate a level of OCC torque capacity (Ton) for achieving adesired output torque profile (TOS, des) during torque phase when OGCremains locked. More specifically, a torque profile τos,des can bespecified to smoothly transition the output shaft torque 120 from a timebefore the torque phase at 73 to a time after the torque phase, therebyeliminating or reducing a torque hole. For a given τin or τe at 114, Eq.(5) specifies a level of OCC torque capacity τon (112) required forachieving the target output torque profile τos,des (120).

Alternatively, for a given on (112), Eq. (5) may be used tosystematically determine a target τe (114) or in required for achievingdesired output shaft torque τos,des (120). Once the target level isdetermined, τe or in can be controlled through engine throttle control,spark timing control, intake and exhaust valve timing control, orthrough an auxiliary torque source such as an electric motor. (Note thatengine torque control is coupled to OCC torque control in Eq. (5)).

The inertia phase begins at 73 in FIG. 4 when OGC is released. OGCtransmits torque only at a non-significant level while OCC carriesenough torque capacity, as shown at 122, to slow down input speed 124 sothat it is closer to shaft #2 speed, as shown at 126. Under thiscondition, both Eq. (3) and Eq. (5) can be reduced to:

$\begin{matrix}{\tau_{on} = {\frac{\tau_{{os},{des}}}{G_{on}}.}} & {{Eq}.\mspace{14mu} (10)}\end{matrix}$

Thus, the output shaft torque τos (120) in the inertia phase isprimarily affected by OCC torque capacity τon (122). According to thepresent invention, Equation (10) is used to provide a target OCC torquecapacity τon, during the inertia phase, that is required to achieve aseamless output shaft torque profile τos,des (120) from the torque phaseto the inertia phase. Ton is a feed-forward term. In addition, there isa feed back as well as an effect of a change in engine torque.

FIG. 4 shows reduced input torque during the inertia phase. This istypically due to engine spark timing control according to a commonpractice in a conventional shift control method, enabling OCC to engagewithin a target shift duration without requiring excessive torquecapacity. The shift event is completed when OCC is securely engaged,thereby coupling input shaft 10 and shaft #2 (24). The engine torquereduction then is removed at 130 and the output shaft torque returns toa level 132, which corresponds to an engine torque level in the highgear configuration.

FIG. 5 shows a control flow chart for the synchronous shift control ofthe present invention when the OGC is slipped during a torque phase. Itdescribes a systematic approach to enable the shift control shown inFIG. 4. As previously stated, one of the advantages of this invention isthe decoupling of OCC control, shown inside the dashed line 136, fromengine control 140 and OGC control 144.

Engine torque can be actively and independently managed at 140 through aclosed loop control to achieve a desired OGC slip speed. OGC torquecapacity is adjusted through either closed loop control or open-loopcontrol of its actuator position or actuator force. During a torquephase, a controller first chooses a desired level of output shaft torque(138). It also chooses desired OGC torque at 143. Then, the controlleruses Equation (3) to self-calibrate the required level of OCC torquecapacity at 146. It adjusts OCC actuator position at 148 or its torquecapacity to realize the desired output shaft torque. The controllerevaluates whether the end of the torque phase is reached at 150 basedupon OGC torque capacity level. If it is not, it repeats the controlloop at 153. It re-estimates the desired output shaft torque at 138 andchooses OGC torque capacity at 143 for the next controller time stepk+1.

The end of the torque phase is reached when OGC torque becomessufficiently small or less than a pre-specified threshold, τthresoff, at150. The controller then releases the OGC clutch 152 and moves to theinertia phase control at 154. Equation (10) is used to determine atarget OCC torque at 154 for a seamless output shaft torque transitionfrom the torque phase to the inertia phase.

FIG. 6 illustrates an alternate control strategy that will achieve theoncoming clutch torque characteristics, the off-going clutch torquecharacteristics and the engine torque characteristics that will avoidoutput shaft torque disturbances previously described. As previouslyindicated, in the strategy of FIG. 5, the output shaft torque that ischosen is used to calculate an oncoming clutch torque as shown at block146 in FIG. 5. Regardless of whether the strategy of FIG. 5 or thestrategy of FIG. 6 is used, the objective is to ensure that the enginetorque will be higher throughout the duration of the torque phase thanthe oncoming clutch torque. The engine speed will remain above theoff-going clutch speed during the torque exchange that occurs during thetorque phase as seen in FIG. 4. This prevents a torque reversal.

In FIG. 6, prior to the start of the torque phase at block 212, theoff-going clutch torque will have decreased to a value that is slightlyless than the input torque. This occurs during the preparatory stage asseen in FIG. 4. A desired output shaft torque then is chosen as shown at213 rather than choosing a desired off-going clutch torque following thestep at 213. As in the case of the FIG. 5 strategy routine, a desiredslip is chosen at 214 as seen in FIG. 6. The value chosen is a valuethat will prevent torque source input speed flare during the torquephase. The slip torque depends upon the rate of change of engine speed(α) as well as engine inertia (I) if an engine is the torque source.

After the desired slip is determined at block 214, a target input torqueis determined at block 215. This input torque (τi,tgt) is a function ofdesired output shaft torque. The target input torque is that torque thatexists for each control loop of a controller until the shift sequencereaches the end of the torque phase. If the sum of the target inputtorque and the desired slip torque is less than a precalibrated maximumvalue, as shown at block 216, the routine will continue to block 218where a change in input torque (Δτi) in any instant during the torquephase is equal to the target input torque (Ti,tgt) minus the change ininput torque (Δτi) at the beginning of the torque phase. If the sum ofthe target input torque and the slipping clutch torque at 216 is greaterthan Ti maximum, the routine is recalculated at 217 until the inquiry at216 is true.

The oncoming clutch target torque (τon,tgt) is computed by determiningthe sum of the delta off-going clutch torque at 219 (change of torque)and the delta input torque calculated at 218 at the end of the torquephase. The input torque then is ramped upwardly to the target. This isthe value for oncoming clutch torque at the end of the torque phase. Thestep of ramping the input torque is shown at 223 in FIG. 6. If theresult of the ramping at 223 is an off-going clutch torque that is lessthan the off-going clutch threshold value, which is precalibrated, theoff-going clutch will be released at shown at 225. As in the case of theroutine of FIG. 5, the routine proceeds through the inertia phase wherethe desired oncoming clutch torque is determined by the equation shownat 226.

The routine 311 of FIG. 7 is somewhat similar to the routine 211 of FIG.6 except that, for example, a desired target oncoming clutch torque ischosen following the start of the torque phase at 312. This is shown atblock 313 in FIG. 7. In contrast, the desired output shaft torque ischosen in the case of FIG. 6 starting at the beginning of the torquephase. After choosing a desired slip at 314, the routine of FIG. 7 willcalculate an input torque at 315 so that the input torque will besufficiently increased to compensate for the target oncoming clutchtorque. This is evident by the rising slope of the input torque plot ofFIG. 4 during the torque phase.

If the target input torque is less than the maximum calibrated inputtorque, as shown at 316, the target input torque and the oncoming clutchtorque target torque are recalibrated at 317 before the routine willcontinue.

If the inquiry at block 316 is true, the routine will advance to block318 where a desired off-going clutch torque is chosen. This is the valueat the end of the torque phase. Having established the desired off-goingclutch torque, the oncoming clutch torque is ramped toward the targetoncoming clutch torque at 319. The clutch actuator for the oncomingclutch torque is adjusted at 321 to achieve the target oncoming clutchtorque. The routine then will continue to block 320 in FIG. 7 where theinput torque is ramped toward the target torque at the end of the torquephase, followed by a controller adjustment at 322 to achieve the target.

A test then is made at 323, as in the case of the routine of FIG. 6, todetermine whether the off-going clutch torque is less than aprecalibrated off-going clutch torque threshold. The threshold torque isdetermined so that a residual torque will be maintained in the clutchactuator rather than having the off-going clutch torque fall to zero.The off-going clutch torque then is released and the routine continuesto the inertia phase as shown at 324 and 325.

It is to be understood that this invention is not limited to the exactshift control steps illustrated and described. Various modifications andequivalents thereof, including revisions to the governing equations (3),(5), (7) and (9), may be made by persons skilled in the art withoutdeparting from the spirit and the scope of the invention to make thisinvention applicable to all types of automatic transmissions, includingboth a lay-shaft type and a planetary type.

While exemplary embodiments are described above, it is not intended thatthese embodiments describe all possible forms of the invention. Rather,the words used in the specification are words of description rather thanlimitation, and it is understood that various changes may be madewithout departing from the spirit and scope of the invention.Additionally, the features of various implementing embodiments may becombined to form further embodiments of the invention.

What is claimed is:
 1. A multiple-ratio transmission wherein apowertrain source provides an input torque to the transmissioncomprising: oncoming and off-going friction elements for effecting ratioshifts; a controller, during a ratio upshift event, configured to:increase the input torque during a torque phase as the off-goingfriction element slips, thereby minimizing torque transients to atransmission output during the upshift event.
 2. The multiple-ratiotransmission set forth in claim 1 wherein the controller is configuredto: decrease the input torque during a transition from the torque phaseto an inertia phase; and increase the input torque at the end of theinertia phase.
 3. The multiple-ratio transmission set forth in claim 2,wherein control of the powertrain source is closed loop control.
 4. Acontrol system for a multiple-ratio vehicle transmission comprising: acontroller for controlling torque capacity of oncoming and off-goingfriction elements, wherein, during an upshift event, torque control of apowertrain source and the off-going friction element is decoupled fromthe oncoming friction element, wherein torque control of the powertrainsource being independently managed to obtain a controlled increase intorque of the powertrain source during an torque phase and prior to aninertia phase.
 5. The control system set forth in claim 4, wherein theoff-going friction element torque capacity is controlled using closedloop control with off-going friction element pressure as a feedbackvariable.
 6. The control system set forth in claim 5 wherein theoff-going friction element torque capacity is open loop controlled. 7.The control system set forth in claim 5 wherein the control systemincludes a clutch actuator wherein a feedback variable of the off-goingfriction element is determined by the position of an actuator for theoff-going friction element.
 8. The control system set forth in claim 4,wherein the controller is configured to control the powertrain sourceand activating pressure of the off-going friction element to achieve adesired slip of the off-going friction element during the torque phase.9. The control system set forth in claim 4, wherein the controller isadapted to determine a desired level of driving torque and torque of thepowertrain source and to self-calibrate a required level of oncomingfriction element torque capacity.
 10. The control system set forth inclaim 9, wherein the controller is adapted to calculate torquetransmitted through the off-going friction element to determine an endof the torque phase prior to the inertia phase.
 11. The control systemset forth in claim 9, wherein the controller is adapted to determine anend of the torque phase at an elapsed time from the beginning of thetorque phase, a corresponding change in torque of the powertrain sourceand a corresponding change in input torque.
 12. The control system setforth in claim 4, wherein the controller is configured to reduce torqueof the powertrain source during the inertia phase and increase torque ofthe powertrain source when the inertia phase ends, whereby input torquechange during the inertia phase is reduced.
 13. A control system as setforth in claim 4 wherein the controller is configured to increase torquecapacity of the oncoming friction element during the torque phase toeffect engagement of the oncoming friction element during the torquephase; and the controller is configured to increase torque of thepowertrain source during the torque phase as torque capacity of theoff-going friction element is reduced during the torque phase to allowslip, whereby the controller manages torque distributed from thepowertrain source.
 14. The control system set forth in claim 13 whereinthe controller is configured to self-calibrate clutch capacity of theoncoming friction element during the torque phase to achieve a desiredinput torque output profile with respect to time before the upshiftevent is completed as the off-going friction element slips during thetorque phase.
 15. The control system set forth in claim 14 wherein thecontroller adjusts torque of the powertrain source to maintain slip ofthe off-going friction element during the torque phase at a desiredlevel.
 16. The control system set forth in claim 15 wherein torque ofthe powertrain source is adjusted during the torque phase with closedloop control using a measured slip of the off-going friction element asa feedback variable.
 17. The control system set forth in claim 14wherein the controller is configured to reduce torque of the powertrainsource during the inertia phase.
 18. The control system set forth inclaim 13 wherein the controller is further configured to establish apreparatory phase prior to the torque phase and to reduce torquecapacity of the first off-going friction element during the preparatoryphase to prepare for its release.
 19. A method for controlling anupshift of a multiple-ratio vehicle transmission comprising: increasingan input torque from a powertrain source during a torque phase;decreasing the input torque during a transition from the torque phase toan inertia phase until the upshift is substantially ended; decreasing anoff-going friction element torque capacity during the torque phase whileallowing slip; and increasing an oncoming friction element torquecapacity during the torque phase.
 20. The method set forth in claim 19further comprising: choosing a desired output shaft torque during thetorque phase; choosing a desired off-going friction element torque;calculating an oncoming friction element torque as a function of thedesired off-going friction element torque and the desired output shafttorque; controlling an oncoming friction element to achieve thecalculated oncoming friction element torque; controlling the powertrainsource to achieve a desired off-going friction element slip; adjustingan off-going friction element pressure to achieve the desired off-goingfriction element torque; and releasing the off-going friction element atthe end of the torque phase.
 21. The method set forth in claim 20further comprising determining whether the desired off-going frictionelement torque is less than a predetermined threshold value beforeallowing the off-going friction element to be released when the torquephase ends.
 22. The method set forth in claim 21 further comprisingdetermining the oncoming friction element torque during the inertiaphase as a function of desired output shaft torque and gear ratio forthe oncoming friction element and for the off-going friction element.23. The method set forth in claim 19 further comprising: choosing adesired transmission output torque; choosing desired off-going frictionelement slip torque during a torque phase of an upshift; calculating atarget input torque as a function of desired transmission output torqueand a delta input torque based on the target input torque; choosing adesired off-going friction element torque; calculating a delta off-goingfriction element torque; calculating oncoming friction element targettorque based on delta off-going friction element torque; rampingoncoming friction element torque to the oncoming friction element targettorque; and ramping the input torque toward the target input torque,whereby a smooth transition is made to an inertia phase as during whichtorque of the powertrain source is decreased during the inertia phase.24. The method set forth in claim 19 further comprising: choosing adesired target oncoming friction element torque during a torque phase ofan upshift; choosing a desired input torque slip; calculating an inputtorque of the powertrain source to compensate for the desired targetoncoming friction element torque; choosing a desired off-going frictionelement torque; ramping the oncoming friction element toque toward thetarget oncoming friction element torque; and ramping the input torquetoward the target torque prior to release of the off-going frictionelement as a transition is made to an inertia phase whereby torquedisturbances in torque delivery through the transmission are reducedduring an upshift.
 25. The method set forth in claim 19 furthercomprising: choosing a desired output shaft torque and a desired slip ofthe off-going friction element; calculating a target input torque fromthe powertrain source as a function of desired output shaft torque;calculating an oncoming friction element torque target, based on adesired off-going friction element torque; and increasing torque of thepowertrain source toward the target input torque followed by release ofthe off-going friction element as a transition is made from the torquephase to an inertia phase in the upshift.
 26. The method set forth inclaim 19 further comprising: choosing a desired oncoming frictionelement torque target; choosing a desired input slip torque; calculatingan input torque target for the powertrain source using the oncomingfriction element torque target and the desired input slip torque;choosing a desired off-going friction element torque; ramping oncomingfriction element torque toward the oncoming friction element target; andramping input torque from the powertrain source toward the target inputtorque and releasing the off-going friction element as the torque of thepowertrain source is decreased at the start of an inertia phase of theupshift.